The present invention relates to a two-stroke cycle gasoline engine, and, more particularly, to a two-stroke cycle gasoline engine adapted for use with automobiles.
A two-stroke cycle engine has theoretically the advantage that an engine of a certain size can generate a greater power than a four-stroke cycle engine of a bigger size because the two-stroke cycle engine has twice as many work cycles per revolution as the four-stroke cycle engine. In fact, however, the conventional two-stroke cycle gasoline engine employing a carburetor has such drawbacks that it has high fuel consumption as compared with the four-stroke cycle engine due to the loss of air-fuel mixture caused by the direct escape, i.e. blow-out, of scavenging mixture to an exhaust manifold during scavenging, and that it cannot generate such a high power as expected from the fact that it has twice as many work strokes as the corresponding four-stroke cycle engine, due to the fact that the scavenging is still insufficient. Because of these problems, the practical use of two-stroke cycle gasoline engines is nowadays limited to the field of small engines which must be simple in structure and low in manufacturing cost.
Conventional two-stroke cycle gasoline engines of the abovementioned type, therefore, generally employ crankcase compression for scavenging. However, the scavenging by crankcase compression is not fully effective and can only provide a relatively low volumetric efficiency. This is the principal cause of the poor output power of conventional two-stroke cycle gasoline engines. In fact, a volumetric efficiency as high as 80% is available in four-stroke cycle engines, while on the other hand the volumetric efficiency of typical two-stroke cycle engines is still as low as 40-50%. The pump stroke volume of crankcase compression is equal to the stroke volume of the engine. However, since the crankcase has a relatively large clearance volume, the compression ratio of crankcase compression is relatively low, so that as a result the amount of air-fuel mixture drawn to the crankcase is small, the amount of delivered mixture is small, the delivery pressure is low and hence the scavenging pressure is low, and consequently it is hard to supply a really adequate amount of scavenging mixture into the power cylinder. As a result, the delivery ratio obtained in an engine wherein scavenging is effected only by the normal crankcase compression is only as high as 0.5-0.8. Since further the trapping efficiency is about 0.7, the volumetric efficiency becomes as low as 40-50% as mentioned above.
The purpose of scavenging is to push the residual exhaust gases in the power cylinder out of it by fresh mixture, and therefore if the pressure of the residual exhaust gases and the distance between the scavenging port and the exhaust port are given, the time required for completing scavenging is determined, provided that stratified scavenging is performed. Now, if the scavenging pressure is low, as when crankcase compression is used, a relatively long time is required for completing scavenging, particularly when the scavenging is performed by uniflow scavenging, and therefore, when the engine is rotating at high speed, it may well occur that the exhaust port is closed before the scavenging is completed so that a large amount of exhaust gases still remains in the power cylinder, and thereby only a very small amount of fresh mixture is charged into the power cylinder. Therefore, conventional two-stroke cycle engines have been unable to operate satisfactorily in the high speed range.
Furthermore, when scavenging depends upon the crankcase compression, since a power piston operates as a pump piston, as a matter of course the operational phase of the scavenging pump means is shifted from that of the power cylinder-piston assembly to which it supplies scavenging mixture by exactly 180.degree., so that when the power piston of the power cylinder-piston assembly reaches its bottom dead center (BDC), the pump piston of the pump cylinder-piston assembly (which is identical with the power piston) simultaneously reaches its top dead center (TDC). In the present description the dead center at the end of compression stroke is called "top dead center" and the opposite dead center is called "bottom dead center". However, as apparent from the crank angle diagram of FIG. 1, the usable scavenging period of the power cylinder-piston assembly extends after BDC of the power piston so that half of it is still left when the power piston reaches its BDC, and in spite of this since at this point the pump piston begins to retreat towards its BDC the pressure in the crankcase greatly lowers, as shown in FIG. 2, so that although the scavenging port is still open the scavenging period is not effectively utilized.
Therefore, it is an object of the present invention not to use crankcase compression for scavenging, but instead to use an independent scavenging pump means having a small clearance volume and a high compression ratio so that the volumetric efficiency of the scavenging pump means is substantially increased and so that the delivery amount of the scavenging pump means is substantially increased, for a fixed total stroke volume of the scavenging pump means. In this connection, another object of the present invention is further to increase the total stroke volume of the scavenging pump means so as to be 1.15-1.65 times as large as the total stroke volume of the power cylinder-piston assembly, so that this, in combination with the use of an independent scavenging pump means, provides substantially increased pressure and amount of scavenging mixture when compared with crankcase compression. In this case, therefore, the power cylinder is scavenged by mixture of no lower pressure and amount than if it is scavenged by mixture supplied by, for instance, a particular crankcase compression system employing a stepped piston and having an increased stroke volume such as 1.35-1.85 times as large as the stroke volume of the power piston. Thus, in accordance with the present invention, high pressure and amount of scavenging mixture necessary for scavenging the power cylinder at high scavenging efficiency are ensured, and if this feature is combined with employing the optimum phase difference between the scavenging pump means and the power means, as explained hereinunder, the volumetric efficiency of the power cylinder can be increased to as much as 75-100%, whereby a two-stroke cycle gasoline engine of the present invention can generate substantially higher power per unit stroke volume of the engine than a conventional two-stroke cycle gasoline engine.
In the present description the total stroke volume of a power cylinder-piston assembly of an engine means the total stroke volume displaced by the power piston when it moves between its bottom dead center and its top dead center, and therefore the effective stroke volume displaced by the power cylinders after the exhaust port is closed by the power piston until they reach TDC is smaller than this aforementioned total stroke volume. Furthermore, when two or more power cylinder-piston assemblies are included in the engine, the total stroke volume of the engine as defined here is the value which is obtained by multiplying the number of power cylinder-piston assemblies by the abovedefined total stroke volume of each power cylinder-piston assembly. The total stroke volume of the pump means the sum of the volume or volumes displaced by the pump piston or pistons while it or they move during their compression stroke.
In connection with the abovementioned objects of the present invention, another object of the present invention is, by the use of an independent scavenging pump means instead of crankcase compression, and by the increase of scavenging pressure, to shorten the time required for scavenging so that the scavenging efficiency is increased up to 80-90% and high power operation of a two-stroke cycle gasoline engine is ensured even in a relatively high-speed operational range.
However, it is to be noted that the relatively high speed operational region contemplated in the present invention means such an operational region in which the conventional, particularly the uniflow scavenging type, two-stroke cycle gasoline engine is unable to operate with sufficient output power, due to insufficient scavenging at high rotational speed. In fact, the aforementioned relatively high-speed rotational region is located in a lower speed region than the high rotational speed region of conventional automobile four-stroke cycle gasoline engines, as explained hereinunder. Therefore, it is still another object of the present invention to provide a two-stroke cycle gasoline engine which can operate in such a lower speed operational region so as to generate sufficient output power. Conventionally, a relatively small-sized four-stroke cycle engine for automobiles is designed so as to be operated at relatively high rotational speed so that relatively high power output is available from a relatively small size engine. In this connection, it is noted that, for example, in the case of an engine which has a two liter piston displacement and produces 92 PS of brake horsepower at 5000 rpm, a large proportion of the power, such as 52 PS out of the indicated power of 144 PS, is consumed by internal friction losses in the engine. The ratio of the internal friction loss to the output power of the engine is substantially reduced by lowering the rotational speed of the engine. In view of this, still another object of the present invention is to utilize the advantage of the fact that a two-stroke cycle engine has twice as many work strokes as a four-stroke cycle engine by increasing the volumetric efficiency of the power cylinder, and to provide an engine which produces high effective power output per unit stroke volume of the engine without increasing the rotational speed to such a high range as in conventional relatively small four-stroke cycle automobile engines. The maximum rotational speed of the engine contemplated in the present invention is 3800 rpm.
As methods of scavenging in two-stroke cycle engines are conventionally known cross scavenging, loop scavenging, and uniflow scavenging. In this connection, and in connection with the aforementioned high pressure scavenging contemplated in the present invention, if the scavenging pressure is increased in cross or loop scavenging, the flow of scavenging mixture is liable to penetrate through the layer of exhaust gases existing in the power cylinder in a short-cutting manner, and also scavenging mixture and exhaust gases may be mixed with each other, thereby not only causing poor scavenging but also increasing the above-explained blow-out loss of mixture, thus lowering the volumetric efficiency. On the other hand, it has been experimentally confirmed that, when uniflow scavenging is employed, it is possible to push the exhaust gases existing in the power cylinder uniformly out of it by the scavenging mixture at high pressure without causing any detrimental mixing between the scavenging mixture and the exhaust gases.
Therefore, it is still another object of the present invention to provide a two-stroke cycle gasoline engine in which high pressure scavenging and uniflow scavenging are combinedly incorporated.
Furthermore, the present invention proposes as an important feature to provide a particular operational phase relation between a power cylinder-piston assembly and a pump cylinder-piston assembly which supplies scavenging mixture to the pump cylinder-piston assembly such that the top dead center of the pump cylinder-piston assembly is, as viewed in the crank angle diagram, in a range between 15.degree. in advance of and 15.degree. behind the midpoint between the bottom dead center and the scavenging port closing point of the power cylinder-piston assembly. This is so determined for the following reasons.
From the viewpoint of increasing the volumetric efficiency of the power cylinder-piston assembly it is desirable that scavenging mixture should be supplied to the power cylinder-piston assembly over the entire scavenging period shown in the crank angle diagram of FIG. 1. However, when for example a power cylinder-piston assembly is supplied with scavenging mixture by a single acting pump cylinder-piston assembly driven by the power cylinder-piston assembly in synchronization therewith with a phase difference, if the TDC of the pump is at the phase position Sc at which the scavenging port of the power cylinder-piston assembly closes, in order to satisfy the aforementioned entire-period scavenging, the crank angle between the BDC of the pump piston and the phase position So at which the scavenging port of the power cylinder-piston assembly opens becomes small. This means that the stroke which the pump piston moves from the beginning of the pump compression stroke until the scavenging port opens is relatively small and that therefore the scavenging pressure which is available when the scavenging port is first opened is relatively low. Therefore, the phase difference between the scavenging port opening phase So and pump BDC must be greater than a predetermined value, so that an object of the present invention, which is to perform scavenging by employing mixture at high pressure, may be accomplished. In accordance with the present invention, when a greater importance is given to the high rotational speed performance of the engine, it is contemplated that the scavenging pressure at the time of opening of the scavenging port (So) should be 0.5-0.6 atm (gauge pressure). The scavenging pressure at the opening of the scavenging port (So) is determined by various factors such as the volumetric efficiency of the pump, compression ratio of the pump at So which is determined by the clearance volume of the pump and the passages extending from the pump to the scavenging port and the length of the pump piston stroke between pump BDC and the point So, and suction and delivery inertia of the pump, etc. When a reed valve is provided in the vicinity of the scavenging port for the purpose of preventing blow-back, the scavenging pressure immediately after the point So is affected by the transient response performance of the reed valve. When the responsiveness of this reed valve is low, i.e. its opening is retarded, the scavenging pressure temporarily lowers after So. Therefore, in consideration of all these factors, crank angle difference between the pump BDC and the scavenging port opening phase point So is determined.
On the other hand, even when the pump piston reaches its TDC at a middle portion of the latter half of the scavenging period, i.e. the period between power piston BDC and scavenging port closing phase point Sc, the scavenging pressure will not in fact immediately lower to zero, as the flow of scavenging mixture is maintained for a certain period after pump TDC, due to the time required for the scavenging mixture to flow from the pump to the scavenging port, the inertia of the scavenging mixture, the retardation effect that the scavenging mixture enters into the power cylinder only after a certain time delay due to the throttling effect applied to the scavenging mixture when it flows through the delivery port of the pump, retardation caused in the flow of scavenging mixture due to delay in response of a reed valve when such a valve is provided, etc. Therefore, even when pump TDC is situated at a middle portion of the latter half of the scavenging period, it can happen that in fact scavenging mixture flows into the power cylinder over the entire region of the scavenging period. In this connection, the time required for the mixture from the scavenging pump to enter into the power cylinder through the scavenging port is determined by the pressure difference across the scavenging port and the throttling ratio of the scavenging port, and this time is not directly concerned with the rotational speed of the engine. On the other hand, the time lapse between pump TDC and the scavenging port closing phase point Sc is shorter as the rotational speed of the engine is higher. Therefore the importance of the crank angle difference between pump TDC and the point Sc varies in accordance with the design of the engine depending to which rotational speed of the engine the most importance is given. From this point of view, therefore, it is not very important that pump TDC should be brought closer to Sc. In consideration of the various abovementioned factors, as a result of experimental researches we have obtained the abovementioned condition with regard to the phase position of pump TDC.
The particular phase position within the aforementioned phase range at which pump TDC is actually positioned is determined in consideration of various factors such as the magnitude of crank angle between So and Sc, the aforementioned factors for determining the scavenging pressure at So, factors for determining the time required for the scavenging mixture to finish flowing into the power cylinder after pump TDC, what rotational speed of the engine is considered most important in the design of the engine, etc. Then an engine is manufactured for experimental tests, and is tested with regard to how the performance of the engine changes in accordance with modifications of various factors and conditions as mentioned above. As a result of such experiments it is possible to determine the particular design of the engine which has the most desirable performance in view of the objects of the present invention.
In the system wherein TDC of a pump piston is at the same phase point as BDC of a power piston the scavenging pressure at the scavenging port opening phase point So becomes very high. Therefore, in such a system, resort is often had to the provision of a mixture tank between the pump and the power cylinder-piston assembly so as to increase the base volume involved in the scavenging system and so as to lower the scavenging pressure at the scavenging port opening phase. When a mixture tank is employed, a reed valve is provided between the pump and the mixture tank, and in the part of the scavenging period after pump TDC the scavenging mixture is only moderately delivered at low pressure from the tank and the effectiveness of the scavenging is reduced. On the other hand, if the clearance volume of the pump is increased (this corresponds to the case wherein no reed valve is provided between the pump and the mixture tank), the compression ratio of the pump lowers, and the suction and delivery performance of the pump also lowers. However, if the scavenging pressure is very high, the scavenging mixture will blow-out to the exhaust manifold through the residual exhaust gases existing in the power cylinder, or the scavenging mixture will mix with the exhaust gases so as also to cause blow-out of scavenging mixture to the exhaust manifold. However, in accordance with the present invention, the scavenging pressure is maintained at a proper level without employing a mixture tank or without providing an additional clearance volume in the pump, thereby effectively utilizing scavenging mixture. In accordance with the present invention, in order to improve pump performance and effective utilization of scavenging mixture, it is rather desirable that the clearance volume should be as small as possible, and that when blow-back from the power cylinder does not occur, and when pump TDC is not very close to power piston BDC, no reed valve should be provided in the scavenging passage.
As mentioned above, in accordance with the present invention, when predominance is given to the performance of the engine in high rotational speed operation, the scavenging pressure is increased up to the order of 0.5-0.6 atm (gauge pressure). Such a high scavenging pressure is not available from conventional crankcase compression. With conventional crankcase compression generally only scavenging pressure of about 0.3-0.35 atm is available, and in the case of an engine having new and particularly improved design scavenging pressure of 0.45 atm is available at the highest. The use of a higher scavenging pressure in the present invention is based upon the recognition that since the time required for the scavenging mixture to flow into the power cylinder and to push the exhaust gases out of the exhaust port is determined by the pressure difference between the scavenging pressure and exhaust pressure and the distance to be travelled by the scavenging mixture while it flows from the scavenging port to the exhaust port, and is not directly concerned with the rotational speed of the engine, if scavenging is to be completed before the exhaust port is closed by the power piston in high speed operation of the engine, the quantity of scavenging mixture must be increased. The volumetric efficiency of a reciprocating piston pump can exceed 100% at a certain rotational speed, if it is properly designed. However, since the rotational speed widely varies in the case of an automobile engine, if the volumetric efficiency of the power cylinder is to be increased up to 75-100% or more, the amount of scavenging mixture must be increased. In view of this, as mentioned above the present invention contemplates to employ an independent pump cylinder-piston assembly having total stroke volume 1.15-1.65 times as large as the total stroke volume of the power cylinder-piston assembly. However, as explained above, even in uniflow scavenging, if the amount of scavenging mixture is too much increased, blow-out of scavenging mixture to the exhaust manifold increases. The ratio of 1.65 proposed by the present invention as the upper limit of the ratio of the total stroke volume of a pump cylinder-piston assembly to that of a power cylinder-piston assembly has been experimentally obtained as the upper limit for avoiding an undesirable degree of blow-out of mixture to the exhaust manifold if the combination of the performance of the change of scavenging pressure and the scavenging period is favorably adjusted.
For example, let us assume that the volumetric efficiency of a pump is 80%, and that 85% of the scavenging mixture delivered from the pump is actually supplied to the power cylinder due to obstruction by a reed valve, etc. Further, let us assume that the mean pressure of the scavenging mixture in the power cylinder is 1.3 atm (absolute pressure). Then, expressing the stroke volume of one power cylinder-piston assembly by Va, and assuming the stroke volume of one pump cylinder-piston assembly which supplies scavenging mixture to said power cylinder-piston assembly to be 1.65 Va, the scavenging volume Vsc of the power cylinder is: EQU Vsc=1.65 Va.times.0.8.times.0.85.times.1/1.3=0.86 Va
Assuming that the power cylinder-piston assembly is of the uniflow opposed piston type, the volume Vec confined in the power cylinder by a pair of pistons when the exhaust side piston closes the exhaust port is expressed by: ##EQU1##
If the second term in the above formula is, for example, 0.30 Va, and if the third term is 0.16 Va (i.e. compression ratio is assumed to be 7.25), EQU Vec=(1-0.30+0.16) Va=0.86 Va
Therefore, in this case: EQU Vec=Vsc,
and this means that scavenging mixture pushes combustion gases completely out of the power cylinder and the scavenging mixture itself is completely retained in the power cylinder with its exhaust port being closed. If there is no leakage of scavenging mixture from the exhaust port, the volumetric efficiency of the power cylinder is 0.86.times.1.3=1.65.times.0.8.times.0.85=1.12, i.e. 112%.
On the other hand, the lower limit of 1.15 with regard to the ratio of the stroke volume of a pump cylinder-piston assembly to the stroke volume of a power cylinder-piston assembly is the value which is considered to be necessary in order to accomplish the objects of the present invention in view of such matters as that, when the engine is small-sized, in some uses a volumetric efficiency of the power cylinder of the order of 75% is acceptable, that high volumetric efficiency of the pump is available by proper design when the engine is normally operated in a relatively narrow range of rotational speed, etc.
Furthermore, currently there exists a great demand for the development of cars which have low fuel consumption, in view of energy saving. Furthermore, cars must satisfy a high standard with regard to the prevention of air pollution. In order to improve fuel consumption, not only the improvement of the fuel consumption of the engine itself but also the reduction of the weight and the air resistance of the vehicle are required. We have noted, in connection with various running tests carried out to prepare for the qualification tests for conforming to the standards for the prevention of air pollution which are becoming more severe nowadays, that fuel consumption is different in summer and in winter due to the difference of atmospheric air density, and we more keenly recognized that the air resistance of the vehicle has an important effect on the fuel consumption of the vehicle even in low speed running. In order to lower the air resistance of the vehicle it is important to reduce the height of the vehicle as much as possible and to form the external shape of the vehicle in a streamlined shape. Particularly it is very effective to lower the engine hood. In order to reduce the height of the vehicle it is effective to eliminate the drive shaft for driving the rear wheels so that the shaft tunnel is eliminated and a flat floor is available, over the entire floor area, thereby constructing a vehicle body having a low floor and a low roof. A method for accomplishing this is to employ the FF system, i.e. the front engine - front drive system. In order to lower the engine hood by a large amount in an automobile employing the FF system while ensuring necessary leg room for the driver and the front seat passenger, it is necessary to reduce substantially the height and length of the engine compartment. Furthermore, in order to reduce the air resistance of the vehicle, it goes without saying that the frontal area of the vehicle must be reduced. Therefore, the width of the vehicle should be minimized. Furthermore, since the transmission, differential gears, and other driving mechanisms must be housed in the engine compartment together with the engine, in the FF system, the space allowed for the engine is much reduced. Light trucks are often designed with the engine mounted under the driver's seat, and in such a design the engine, being relatively long, often extends so far backward as to make a hump of the engine enclosure rearward of the cabin, thus shortening the deck.
It is therefore still another object of the present invention to deal with the aforementioned problems and requirements and to provide a small size gasoline engine having a low height, a small length and not a very large width, yet being capable of generating high power.
As uniflow scavenging engines are known an engine having horizontally opposed pistons, an engine having an exhaust poppet valve, etc. In order to accomplish the aforementioned objects of the present invention, we now consider an engine having horizontally opposed pistons. That is, it is found that an engine having a power cylinder-piston assembly employing horizontally opposed pistons is particularly advantageous.
Therefore, in order to accomplish the aforementioned objects, the present invention proposes to employ at least one two-stroke cycle power cylinder-piston assembly incorporating uniflow scavenging and two horizontally opposed pistons as the power cylinder-piston assembly of the engine. By combining such a power cylinder-piston assembly with the aforementioned concept of high flow and pressure of scavenging mixture and also with the particular phase condition, it is possible to charge the power cylinder with fresh mixture with high volumetric efficiency without causing substantial blow-out of scavenging mixture to the exhaust manifold, and because of this, it is possible to obtain an engine of reduced height and length having the high power generating ability even at relatively low rotational speed, when compared with a conventional four-stroke cycle engine. Furthermore, in contrast to the emission performance of the conventional two-stroke cycle gasoline engine, which shows high concentration levels of HC in the exhaust gases, such as 5-10 times as high as those of the conventional four-stroke cycle gasoline engine, the engine of the present invention is able, due to substantial avoidance of blow-out of scavenging mixture to the exhaust manifold, to keep HC concentration in the exhaust gases at a sufficiently low level.
In connection with the aforementioned concept of employing at least one two-stroke cycle power cylinder-piston assembly incorporating uniflow scavenging and two horizontally opposed pistons as the power cylinder-piston means of the engine, the present invention further proposes to employ at least one pump cylinder-piston assembly of the reciprocating type as the scavenging pump means. By employing such a pump cylinder-piston assembly it is possible to ensure the necessary amount and pressure of scavenging mixture even in low speed operation and it is also possible to construct the scavenging pump with a simpler and less expensive structure.
When compared with this, if a rotary pump is employed, although the advantage that scavenging mixture is supplied to the scavenging ports throughout the entire scavenging region is obtained on the one hand, on the other hand scavenging pressure is constantly applied to the scavenging port even during the non-scavenging period, whereby it may happen that scavenging mixture should leak through the clearance between the power cylinder and the piston, thereby increasing the pumping loss, thereby causing a great disadvantage in the case of a small-sized engine. In contrast, when a reciprocating piston pump is employed, the operational phase of the piston can be properly matched to the operational phase of a power cylinder-piston assembly so that the required scavenging pressure is generated only when it is required by the power cylinder-piston assembly.
In this connection, furthermore, if a pump cylinder-piston assembly incorporating horizontally opposed pistons is employed as a reciprocating pump in combination with the aforementioned two-stroke cycle power cylinder piston assembly incorporating uniflow scavenging and horizontally opposed pistons, another advantage is obtained in that more desirable harmony between the dimensions of the power cylinder-piston assembly and of the pump cylinder-piston assembly is available.
In more detail, a two-stroke cycle power cylinder-piston assembly incorporating uniflow scavenging and horizontally opposed pistons has a volume to be scavenged slightly more than twice as much as the stroke of the individual pistons. Therefore, if the power cylinder must be scavenged by a scavenging pump having a single piston, either the diameter of the pump cylinder or the stroke of the pump piston must be relatively large. In either case, in view of the fact that the total stroke volume of the scavenging pump means is to be 1.15-1.65 times as large as the total stroke volume of the power cylinder-piston assembly, particularly since crankcase compression is not employed, it is apprehended that either the width or the length of the scavenging pump means may become too large compared with those of the power cylinder-piston assembly. However, if the scavenging pump means is provided as a pump cylinder-piston assembly having horizontally opposed pistons, it is possible to maintain both the diameter of the pump cylinder and the stroke of the pump piston within reasonable values so as to provide desirable harmony with the power cylinder-piston assembly. When such a pump cylinder-piston assembly is arranged horizontally side by side with a power cylinder piston assembly of the same type having horizontally opposed pistons, the engine presents a compact overall configuration like a horizontally flat block, rectangular in a plan view. An engine for a small size or light automobile will comprise, at the most, one or two two-stroke cycle power cylinder-piston assemblies of the aforementioned type incorporating uniflow scavenging and horizontally opposed pistons. In this case the balancing of the scavenging pump is important. If one power cylinder-piston assembly incorporating uniflow scavenging and horizontally opposed pistons is served by a single cylinder-single piston scavenging pump, the pump piston will become relatively large, requiring a relatively large counterweight, resulting in a relatively large crankcase, yet perfect balancing of reciprocating masses will not be attained. However, if the pump is a cylinder-piston assembly having horizontally opposed pistons, inertial forces of the reciprocating masses related to individual opposed pistons are perfectly balanced, whereby the crankcases for individual pistons are substantially reduced in size together with reduction of the height and length of the engine, thereby providing a compact two-stroke cycle engine of the horizontally opposed piston type less prone to vibration.
However, the differences of engine volume and of dynamic balance between a single piston scavenging pump and an opposed piston scavenging pump will become less important as the engine becomes smaller, while on the other hand if the engine becomes smaller, the difference in manufacturing cost, which is governed by structural complexity, will become more important. Therefore, it must be individually decided according to various conditions which of the two factors should have priority over the other. When a pump cylinder-piston assembly having horizontally opposed pistons is employed as the scavenging pump means, the reciprocating inertia forces in the pump means are well balanced, and this, in combination with a power cylinder-piston assembly of the same horizontally opposed piston type in which the reciprocating inertia forces are also well balanced, can provide a well balanced, less prone to vibration, and quiet engine.
With respect to a pair of crankshafts of the power and pump cylinder-piston assemblies of the horizontally opposed piston type, if they are rotated in opposite directions, moments produced by forces perpendicular to the crankshafts are also balanced. However, this requires incorporating a rotation reversing mechanism including an idle gear between the two crankshafts, and therefore increases manufacturing cost. Therefore, as an embodiment of the present invention, it is proposed to connect drivingly a pair of crankshafts of the power and pump cylinder-piston assemblies of the horizontally opposed piston type simply by an endless chain so that the two crankshafts rotate in the same direction. In this regard, it is a matter of choice between pursuing quietness of vibration in engine operation and pursuing reduction of cost to select the system of mutual counterrotation of a pair of crankshafts or to select the system of rotation in the same direction of a pair of crankshafts, and this is, in any event, a matter of design with regard to the engine of the present invention.
When the two-stroke cycle gasoline engine of the present invention comprises, for example, two two-stroke cycle power cylinder-piston assemblies, since the crankcases of these power cylinder-piston assemblies are not utilized for crankcase compression of scavenging mixture, the scavenging pump means to serve for the two power cylinder-piston assemblies must have a relatively large capacity. Therefore, even when the scavenging pump is constructed as a pump cylinder-piston assembly having horizontally opposed pistons, a single acting pump cylinder-piston assembly of the horizontally opposed piston type will not be sufficient to supply the necessary flow of scavenging mixture. Furthermore, when two power cylinder-piston assemblies are combined to operate with phase difference of 180.degree. therebetween, another difficulty is encountered with regard to the matching of the operational phase of the scavenging pump to that of the power cylinder-piston assemblies. In view of these problems, the present invention further proposes to employ a double acting pump cylinder-piston assembly having horizontally opposed pistons so as to make the two actions of the pump pistons serve for the scavenging of the first and second power cylinder-piston assemblies, respectively. By this arrangement, it is possible to supply scavenging mixture to two power cylinder-piston assemblies by using one pump cylinder-piston assembly while maintaining harmony between the dimensions of the power cylinder-piston assemblies and of the pump cylinder-piston assembly, and thus an engine having high power output relative to its volume is obtained.